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TROUBLESHOOTING CENTRIFUGAL COMPRESSOR OIL SEALS
AUDITING THE EXISTING OIL SEAL DESIGN
DURING COMPRESSOR RERATING
TEN CRITICAL STEPS TO A SUCCESSFUL AUDIT
Abstract
Background
Oil seal technology remains the primary sealing
technology for existing compressors. However,
vendor support for seal oil systems is waning. As
production demands increase, existing seal oil
systems will require revamp and reapplication
unless Dry Gas Seal technology can be justified. A
step increase in seal oil pressure and resulting flow
without proper reengineering of the system
components can result in significant operating
problems.
Centrifugal compressors utilize shaft end seals of
three basic designs. Dry Gas Seals, Oil Face Seals
and Oil Bushing Seals.
This paper will cover:
•
Basic technology of Compressor
Components and Seal Oil System
Design
•
Audit considerations during
compressor revamp and reapplication
of seal oil systems
•
Warnings, Checkpoints
and a Case History
Of these, the majority of older existing installations
utilize Oil Seals. As these compressors are
evaluated for process flow and pressure uprates,
the shaft end seals must be audited to assure
flawless operation at the new operating conditions,
as well as assure safety to operating personnel.
Often, limited turnaround time and price for dry
gas seal retrofits, dictates that existing oil seal
technology must be utilized. As support personnel
at the O.E.M. (Original Equipment Manufacturer)
for these types of seals are reduced, a detailed
checklist for auditing the oil seals as well as their
support ancillaries is required.
Audit
During rerating of centrifugal compressors with oil
type seals, performing the following ten key steps,
is critical in assuring flawless execution of the
uprate commissioning and long term reliability of
the equipment.
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STEP ONE - Determine the complete range of revised compressor
operating pressures
Figure one is typical gas compression P&I diagram.
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Revised Operating Pressures, Total Operating Range
The seal oil system, supporting the shaft end seals must operate over a wide range of oil pressures and
resulting seal oil flow rates. These include operation at the following conditions:
WARNING:
•
The pressurization rate must be slow enough for the Seal Oil System to track the gas pressure
spike and maintain the required seal oil to gas differential. If the pressurization rate is too fast
the compressor will blow gas into the process area. An orifice may be required in the loading line
to reduce the loading rate.
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STEP TWO - Determine the revised compressor seal gas reference
pressures
Once all the process operating conditions are defined, the revised compressor seal oil operating pressures
must then be determined. This is a function of the seal type, the compressor gas reference design, and the
seal control type.
Seal Type
Centrifugal compressor oil seals are grouped into two categories. Oil face type and oil bushing type.
Oil Face Type Seals
Figure three is an example of an oil face type seal. This type of seal utilizes a highly lapped carbon ring that
runs against a rotating ring. The operating oil pressures of this seal is typically controlled at 35 - 50 psid.,
above the compressor gas reference pressure. Contaminated seal oil leakage rates are dependant on seal
diameter, and are measured in gallons per day. (A typical 4 inch seal has a contaminated leakage rate on
the order of 5-7 gallons per day.)
Conventional, self operated differential pressure control valves or a combination of differential pressure
transmitters and receiver controllers, when a pneumatic control valve is utilized, control the 35 - 50 seal oil
to gas differential pressure.
Restrictions on this seal include both speed and maximum operating pressure. The rubbing velocity of the
rotating ring restricts the speed to (approx. 350 ft/min). Maximum pressure tops out at approximately (750
psig).
Figure Three
Oil Face Seal
WARNING:
•
•
Face seals have a speed limitation where the natural frequency of stationary components has
lined up with running speed. Verify that the seal supplier has experience with your seal diameter
at the new revamp design speed. Premature failures of seal components due to resonance have
been experienced.
High speed designs are available, however they require change out of certain seal stationary
components to avoid the resonance with operating speed.
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When the limits of operating speed and pressure for the face type seal are exceeded, oil bushing type seals
are required.
Oil Bushing Type Seal
Figure Four
Oil Bushing Seal
This type of oil seal utilizes a close clearance bushing on the gas side, as opposed to the carbon ring utilized
on the face type design. The minimum operating clearance of this bushing is always limited to a percentage
of journal bearing clearance. To minimize the contaminated leakage rates of this type of seal, the design oil
supply pressure (oil in), is controlled at a lower differential than the face type seal. An overhead seal oil
tank, mounted above the compressor centerline is required to control the seal oil to reference gas
differential to a level of 5-7 psid.
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The following figure shows the overhead tank arrangement for a back pressure control design.
Contaminated seal oil leakage rates are dependent on seal size but are measured in gallons per hour, versus
gallons per day for the face seal. (For a 4 inch seal leakage rates on the order of 3-5 gallons per hour can
be expected. Developments including steps and windback grooving of the seal bores of the gas side bushing
have reduced the contaminated leakage rate, (pumping bushing, Figure Four). These should be considered
during revamps to reduce the contaminated leakage rates and their effects on critical seal oil properties
such as oil flash point when the oil is recycled back into the oil system).
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Compressor gas reference design
Figure 6 depicts a theoretical compressor design utilizing oil face seals, operating at a 45 psi differential
pressure above the gas reference pressures.
This design would require two separate seal oil control stations, one, supplying oil at 145 psig and the other
at 345 psig.
Next a balance piston is added to the design to reduce the axial forces on the rotor, and allow for a smaller
thrust bearing design, figure seven. The model again depicts a face seal, with the gas reference on the
balance piston side of the compressor. This design requires only one seal oil supply control station, as the
gas reference pressures are now "equalized". The slight loss in gas pressure through the balance piston
equalization line, results in slightly larger seal oil to gas differential pressure on the suction end seal, (148 100 = 48 psid).
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The figure eight design is of a higher pressure or speed, requiring the use of a bushing type seal. For these
types of compressors, the impellers are typically arranged in a back to back configuration, to again reduce
the axial forces on the rotor, and minimize thrust. For this design, two overhead seal oil tanks are required,
each with it's own seal oil differential pressure control valve, holding a 5 psi., seal oil to gas differential
pressure.
Next is an equalized seal design which utilizes a multi tooth labyrinth, in an attempt to equalize the gas
reference pressures on each end of the compressor, figure nine. For this design, only one overhead seal oil
tank would be provided, again sensing gas reference pressure on the discharge end of the compressor. More
significantly, is that the slight loss in pressure through the seal equalization line results in significantly
higher seal oil to gas reference pressure on the suction end seal. This higher differential results in higher
contaminated seal oil leakage rates on the suction end seal.
Discharge seal (508 - 503 = 5 psid.)
Suction seal (508 - 500 = 8 psid.)
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CHECKPOINT:
Verify that the revamp balance piston or seal equalization line flow does not result in an excessive pressure
drop; (limit 1-2 psid). Excessive line loss results in a higher level of seal unequalization. The result is a
higher suction end seal, contaminated leakage rates.
CHECKPOINT:
As labyrinth designs improve, (abradable type), the seal clearance can be reduced. The tap connection for
the gas reference can also be designed to reference various gas pressure zones within the seal. Often the gas
reference is taken off the discharge end contaminated seal oil drain connection. This can unfortunately
effect the response of the overhead tank to pressure spikes in the compressor suction. The volume of gas in
the overhead tank can also affect this. The only way for the seal oil pressure to change is to change the gas
pressure in the overhead tank. If a close clearance equalization seal is restricting the flow of gas to the
overhead tank, the result is that the overhead tank will not respond to the gas pressure spikes. The result is
blowing process gas into the compressor deck area during process pressure spikes. A bypass gas reference,
similar to the design shown below may be required.
Figure Ten
Equalization Seal Bypass
Bushing Seal
Backpressure Control
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One final design often encountered in the field is the single stage compressor, typical for pipeline
applications. Here the gas reference pressure, at the back of the impeller is approximately 60% of the total
pressure rise across the compressor. For this example, the design compressor discharge, at the discharge
flange would be 867 psig, with 800 psig at the impeller hub.
STEP THREE - Determine the compressor seal control type
Oil Seal design can utilize Forward Pressure Control, or Back Pressure Control. The control scheme effects
not only the seal design, but also the design and location of the seal oil system ancillaries.
Design A, Face Seal - Forward Pressure Control
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Contaminated seal oil leakage rates across the carbon ring are measured in gallons per day. However the
required cooling flow is in gallons per minute. For this design, the cooling flow is controlled by the
clearance of the breakdown bushing, as well as the size of the bypass orifice.
CHECKPOINT:
Verify the new required cooling flow for the revamp conditions and adjust the bypass orifice size or
breakdown bushing clearance as required. Solenoid operated bypass valves, or variable orifices have
occasionally been utilized to increase oil cooling flows during off design operation.
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CHECKPOINT:
This particular face seal design utilizes a shutdown piston device to keep the seal faces tight, thus
avoiding blowing gas during loss of seal oil. Operation of the individual seal design during
Emergency Blackout Conditions must be reviewed with the O.E.M. If a shutdown device is not
included in the seal design, an alternate oil supply source during power blackout must be considered.
This can include steam, battery, pneumatic or expander driven emergency seal oil pumps. A
pressurized accumulator, similar to the following design, may also be considered.
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Control Scheme
This design utilizes a forward pressure control scheme as shown in Figure Fifteen. The 35-50 psid. seal oil
to gas differential is controlled either by self operating differential pressure control valves, as shown in the
figure, or a differential pressure transmitter / pneumatic controller. The location of the (TEE) connection
determines where the differential pressure is being controlled.
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Forward Pressure Control Valve Flow Summary
Design B, Face Seal - Back Pressure Control
This seal design is different from the forward pressure control, in that the bypass orifice feature is replaced
with a cross flow port. If additional cooling is required for the revamp condition, it is added by increasing
the flow through this port, (oil out).
Figure Seventeen
Face Seal Backpressure Control
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Control for this type of design puts the differential pressure controller on the outlet side of the seal oil flow,
(oil out). A flow control valve is typically added to the supply side, to adjust the total inlet flow to the
required amount. Again the location of the (TEE), connection establishes the exact location where the 3550 psid is being controlled.
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Back Pressure Control Valve Flow Summary
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Design C, Bushing Seal - Forward Pressure Control
For this seal design, Figure Twenty, a close clearance bushing is substituted for the carbon ring.
This design is setup for forward pressure control, which requires use of an overhead seal oil tank.
Contaminated seal oil flows to the drain at a rate in gallons per hour. The remaining oil flow passes across
the atmospheric bushings at a rate measured in gallons per minute.
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A schematic of the overhead tank used to control the 5-7 psid seal oil to gas reference pressure is shown in
Figure Twenty One.
The location of this (TEE) connection is more critical than the examples utilizing the face seals. This is
mainly due to the lower differential pressure control level (5-7 psid) versus the (35-50 psid) for the face
seal. The 5 psi pressure can quickly erode, in the form of piping losses as the new, higher seal oil flow rates
pass through the field piping, compressor O.E.M. supplied piping, and compressor endwall drillings.
WARNING:
•
•
Excessive piping pressure drop, from the (TEE) connection to the compressor endwall
connections can result in the inability to properly seal the compressor. The end result could be
blowing gas during normal operating as well as upset conditions.
Excessive oil "system" pressure drops can result in loss of process gas containment during
emergency, blackout conditions.
(This situation deteriorates as the oil pressure / seal oil flow increase)
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Design D, Bushing Seal - Backpressure Pressure Control
This design utilizes the cross flow, (Oil out) design as depicted on Figure Twenty Two.
Figure Twenty Two
Bushing Seal
Back Pressure Control
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The seal oil overhead tank is located on the outlet side as shown on Figure Twenty Three.
This example shows a cascade design for the seal back pressure valves. At start-up, with zero gas pressure
in the unit, both valves are ramped wide open to avoid flooding the overhead tank, and the discharge end
contaminated seal oil drainer. The location of the (Tee) connection is also important for this design as it
directly effects the operating seal oil to gas differential at the seal face.
WARNING:
Excessive piping pressure drop from the (Tee) connection to the compressor endwall can result in:
• Higher contaminated seal oil leakage rates as a result of higher seal oil to gas differential at the
seal, during normal operation.
• Possible loss of process gas containment during emergency, blackout conditions.
CHECKPOINT:
This type of seal design, combined with the backpressure control scheme, and larger seal diameters,
has lead to subsynchronous vibration problems in the field. The root cause was traced to extending
the seal diameter into previous untested regions. The resulting design acted as a pump as speed
increased. The higher discharge pressure realized at the seal discharge was sensed by the overhead
tank, thus lowering the seal oil supply pressure. The reduced supply pressure unseated the gas side
seal, ( Pumping Bushing In Figure 22), resulting in the subsynchronous vibration. The seal supplier
must verify operating experience at the new operating speed, in order to clear this concern.
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STEP FOUR - Determine the Revised Settleout Pressure
Often the compressor discharge pressure is increased for the new process conditions. This requires resetting
of the discharge separator relief valve. As the majority of oil seals are referenced to compressor suction
pressure, for this example, there will be no change to the design oil seal supply pressure.
Often the Compressor O.E.M. (Original Equipment Manufacturer), will not flag the fact that the settleout
pressure for the above example, will increase. This is typically because this calculation is out of the
O.E.M.'s scope of supply, and within the responsibility of the process designers for the rerate.
The compressor operating pressure, as well as the volume of gas in the piping, coolers and separators,
determines settleout pressure. Often, during major process upgrades, the relief valve set pressure on the
suction separator is increased
During a normal shutdown of the compressor, the driver is tripped, the recycle valve is immediately
opened, to avoid surging the compressor. As the unit coasts down to zero RPM, the compressor suction
pressure rapidly increases until all flow stops and the compressor suction and discharge pressures are
equalized. This pressure is identified as the compressor settleout pressure.
At this point the compressor seal oil system must supply the required oil flow and pressure, to properly seal
the compressor shaft end. Inability to do so will result in process gas blow by thru the seals and into the
compressor deck. This can lead to serious consequences to the plant as well as the operating personnel.
CHECKPOINT:
Settleout pressure is typically limited to the setting of the relief valve, on the suction separator. The set
pressure of this R.V. , plus 10% for full accumulation, is the maximum suction pressure the compressor
will experience. (Assuming this relief valve is designed for full block in flow, and not a fire only type.)
The new P&I should be carefully reviewed to assure the new maximum settleout pressure is properly
calculated.
STEP FIVE - Determine the Revised Seal Oil Flows at all the new
operating conditions
This step requires coordination with the compressor seal O.E.M. Once the new operating conditions are
established, the resulting oil flows must be calculated. The best analogy is to consider the seal as an
equivalent orifice. A change in pressure differential across the equivalent orifice results in a change in oil
flow.
WARNING:
Oil seals are susceptible to overheating and damage, when operated for extended periods of time at
oil supply pressures significantly below design pressure. High pressure applications are more
sensitive to this concern. The Seal Supplier should identify the minimum oil supply pressure versus
speed, for the revised design.
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STEP SIX - Audit the Existing Seal Oil Console
Now that the new seal oil pressures and resulting seal oil flows are fully defined, the seal oil system must
be carefully audited, to assure it can handle all the new operating conditions:
•
•
•
•
•
Operation with zero gas pressure in the unit
Pressurization to suction pressure
Ramp-up to minimum governor
Normal shutdown to settleout pressure
Full emergency blackout
The following areas should be audited, using API 614 :
Lubrication, Shaft-Sealing and Control-Oil Systems and Auxiliaries for Petrochemical, Chemical,
and Gas Industries, as a design basis.
Verify the original oil system hydro pressure, and review the pressure rating of each individual componens
to assure it will be in compliance with local code requirements, at the new operating pressures.
1.
Oil Reservoir
•
•
•
•
•
2.
Verify oil retention time, and working capacity to avoid foaming.
Verify rundown capacity especially if larger overhead seal oil tanks are added.
Verify tank is properly vented, (min. is sum of all return lines), to avoid overpressurization.
Verify pump suction velocity is less than 5 ft/sec., to avoid cavitation.
Verify that the suction strainers will not result in excessive pressure drop.
Seal oil Pumps
•
•
Verify pumps are sized to deliver the revamp system flow plus the greater of either 20 gpm or a 20
percent excess of system flow.
Verify pump will delivery the new required pressure.
Typical pump pressure requirements take into account the following:
(Forward Pressure Control Design)
+
+
+
+
+
+
+
+
•
x
5
10
10
20
5
5
20
10
(x +85)
-
Oil Pressure at Seal Oil Control Point
End Wall & Compressor Piping Pressure Drops
Static Head Loss (Compressor to Console) (This assumes 27 feet of elevation)
Control Valve to Compressor Piping Loss
Control Valve Pressure Drop
Oil Console Piping Pressure Drop
Transfer Valve Pressure Drop
Filter Pressure Drop (Dirty)
Cooler Pressure Drop
Pressure Required at Pump Discharge (Maximum)
Verify pump driver horsepower at cold start and normal operating condtions
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3.
4.
Seal oil Pump Relief Valves
•
Verify relief valves will pass full pump flow at block in.
•
Verify full accumulation pressure does not exceed vessel pressure limits.
•
Verify pump flow can accommodate hydraulic type relief valve parasitic leakage.
Coolers
•
5.
6.
Verify new heat load
Filters / Transfer Valves
•
Verify filter differential at cold start does not exceed cartridge collapse pressure.
•
Verify normal operation differential.
Oil system backpressure control valve
•
Verify valve Cv during all operating conditions does not exceed the following limits:
-
Self operated valve:
25 - 75% open
-
Pneumatic operated valve:
10 - 90 % open
-
Normal operating position:
40 - 60 % open
Governing Equation
Q
=
Cv √ ∆ P/G
Q
=
Flow in Gallons per Minute
Cv
=
Valve Sizing Coefficient
(Number of US gallons of water that will flow through the
valve in one minute at a pressure differential of one)
∆P
=
Valve Pressure Differential
G
=
Specific Gravity of Fluid
(.87 for ISO 32 oil)
•
Verify valve will handle 2 pump running condition
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7.
Seal Oil Differential Pressure control valve
•
Verify Valve Cv is within the above limits for all conditions:
-
8.
Commissioning - zero gas in compressor
Pressurize to suction pressure
Normal shutdown to settleout pressure
Emergency blackout operation
Verify system inner connecting piping is within O.E.M. guidelines.
Figure Twenty Four is a typical guideline for a centrifugal compressor utilizing a back pressure control bushing type
seal. This example depicts both the overhead seal oil tank and a gravity lube oil rundown tank. Both must supply oil to
the compressor during the Emergency Blackout Condition, where all power is lost to the oil pumps.
Oil piping pressure drops from the tanks to the compressor inlet flanges must be within the O.E.M. guidelines to provide
lube oil to the bearings as well as seal oil to avoid blowing gas into the process area.
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STEP SEVEN - Calculate New Process Vent Time
During this period, both suction and discharge valves immediately close, and the vent valve opens. The
vent time is a function of the volume of gas left in the system and the capacity, (Cg), of the vent valve.
The process people should be contacted to calculate the vent time.
WARNING:
The depressurization rate of the system must take into account the limitations of the compressor
and seal O-ring material. This is particularly sensitive on high pressure applications where
O-rings exposed to the process gas can experience failures due to explosive decompression.
Gas pockets trapped in the O-ring material rapidly expand during this period, damaging the
O-ring. Typically the vent rate is limited to approximately 5 psi /sec., however this number must
be confirmed with the oil seal O.E.M. as O-ring material upgrades as well as a change in O-ring
hardness (Durometer Rating) may be required to meet this rate. Additionally, as the operating
pressures move above 1000 psi., the vent rate is often reduced to approximately 2 psi / sec.
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STEP EIGHT: Calculate Oil Volume available for Emergency
Blackout Condition
The first calculation involves the time it takes to vent the process system to zero. This time then is
multiplied by the seal flow to determine the required volume of rundown capacity required to protect the
compressor deck from the blowing of gas during the Power Blackout Condition.
Oil Face Seals
If the seal incorporates a shutdown piston type device, then, no further action is required.
For other types of Face Seal Design, the site must weigh the risk of gas blow by that will occur during loss
of seal oil, and execute the required corrective action. If the power source for the auxiliary seal oil pump is
not stable, an alternate seal oil pump drive or pressurized rundown tanks may be required.
Oil Bushing Seals
As reviewed earlier, this design requires on overhead seal oil tank.
Figure Twenty-Six
API Overhead Seal Oil Tank
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The volume of oil in the tank after trip must be checked to verify it will supply the oil seals with the
amount of flow required to avoid blowing gas into the compressor deck area. If the vent ramp rate cannot
be further increased the possible solutions involve:
- Increasing size of overhead seal oil tanks
- Attempting to redistribute the levels in the overhead tank as defined in API 614.
- Providing a pressurized rundown tank at grade level as shown in Figure Fourteen.
- Considering reduction in breakdown bushing clearance, to reduce flow requirements.
WARNING:
•
•
A reduction in breakdown bushing clearance can result in subsynchronous vibration. A complete
audit of the rotor dynamics must be conducted if this type of modification is being considered.
Any oil volume increase, such as the size of the overhead seal oil tank, requires an audit of the
seal oil reservoir sizing to avoid overflowing during rundown.
CHECKPOINT:
Revamps, involving an increase in operating pressure should consider auditing the rotor dynamics of
the compressor. Oil seals are sensitive to lockup due to excessive axial forces that result when oil
pressures are increased. During high pressure "Hard Starts" of units with oil seals, the atmospheric
breakdown bushings can be eccentric and locked into this position as the unit is ramped up in speed.
This excessive eccentricity can result in vibration problems that may require redesign. A temporary
solution may involve attempting to accelerate the unit with a lower seal oil supply pressure, thus
reducing the axial forces on the bushings, resulting in lower seal eccentricity at speed.
STEP NINE: Calculate Compressor Endwall and Seal Component
Pressure Drop on Oil Supply Side
As discussed earlier, this is most critical on Bushing Type Seals as the seal oil to gas differential is only 5-7
psid.
Forward Pressure Control
For this design, excessive endwall pressure drop can result in loss of containment during maximum flow,
(settleout), or Emergency Blackout conditions.
The endwall drilling, as well as a complete audit of oil porting of the seal components, must be performed.
CHECKPOINT:
Pressure drop, for forward pressure control design should not exceed 1-2 psid., at the maximum flow
condition.
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Back Pressure Control
Excessive endwall pressure drops for this design may result in:
•
High seal oil to gas differential during normal operation, resulting in excessive contaminated leakage
rates.
•
Lack of gas containment during Emergency Blackout Condition
Corrective Action:
1.
2.
3.
Excessive pressure drop can be reduced by reducing the clearance of the breakdown
bushings thus reducing oil flow. Again, the warning about subsynchronous vibration
and rotor response analysis applies.
The existing endwall drilling may have to be increased. This may prove difficult if stub
out piping is also too small. Casing rehydro may also be required.
In certain cases, a forward pressure control design may be able to be converted to a
back pressure design to eliminate the problem. The O.E.M. drawings will have to be
reviewed to determine if this additional porting is available or can be added.
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STEP TEN: Calculate Seal Component and Compressor Endwall
Pressure Drop on Oil Drain Side
WARNING:
The inability to drain the upgrade seal oil flow out of the compressor can result in the flooding of the
bearing and seal housing vents, leading to a possible fire condition.
All seal components, including the endwall drain porting and external drain piping must be audited to
assure they can properly drain the new seal oil flow. The flow velocity should be in the 1-2 ft./sec. range as
a maximum.
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CASE HISTORY
A refinery upgrade required the reapplication of an existing barrel type centrifugal compressor that utilized
oil bushing type seals, forward pressure control. The existing system utilized a combined lube and seal oil
system, utilizing a main oil pump for the lube oil, and a booster pump for the seal oil.
The turnaround was scheduled for 18 days to hydrocarbon in, which with block in and compressor
performance testing, resulted in 13 days of available time for maintenance. This window eliminated the
possibility of attempting to reengineer the compressor for Dry Gas Seals.
Step One - Determine the revised Compressor Operating Pressures
•
•
Original design pressure was 550 psig.
The revised operating pressure was 733 psig.
Step Two - Determine the revised Compressor Seal Gas Reference Pressures
•
The Compressor utilized an equalized seal design, referenced off the discharge end
of the compressor. Therefore only one overhead seal oil tank was utilized.
Step Three - Determine the Compressor Seal Control Type.
•
The Compressor utilized bushing type seals, forward pressure control
Step Four - Determine the Revised Settleout Pressure.
•
•
Original design settleout pressure was 600psig.
The revised settleout was 801 psig
Step Five - Determine the Revised Seal Oil Flows at all the new operating conditions.
•
•
Original design flows were 7 GPM normal and 10 GPM shutdown.
The revised flows were:
NORMAL
•
•
Minimum breakdown seal clearance : 19 GPM @ 733 psig.
Maximum breakdown seal clearance : 26 GPM @ 733 psig.
•
Maximum breakdown seal clearance : 28 GPM @ 801 psig.
SETTLEOUT
Step Six - Audit the Existing Seal Oil Console
•
Reservoir
1.
2.
The working capacity and retention time were less than API requirements.
The return line flow velocity would increase above 5 ft. / sec.
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SAUDI ARAMCO CONSULTING SERVICES
ROTATING EQUIPMENT / VIBRATION TECHNICAL EXCHANGE
•
Pumps
1.
2.
•
Pump Relief Valves
1.
•
Excessive additional pressure drop was found in the oil console piping as well as
the inner connecting piping. This forced the relief valve set point over 95 psig
above the current setting
Seal Oil Filters
1.
•
The existing seal oil pump could not handle the revised shutdown flow. A
proposal was made to increase the seal oil pump size, however this required
operation of both lube oil pumps, when both new seal oil pumps were running.
Pump suction piping velocity would increase to over 10 ft. / sec.
The existing seal oil filters would experience high differential pressures at the
new flows.
Control Valves
1.
2.
The seal oil system backpressure control valve was too small for the 2 pump
operation, if a larger seal oil pump were added.
The seal oil supply control valve was too small for the new seal oil flows.
CONCLUSION: Oil System Audit
Taking into account the turnaround time, and the number of problems encountered
during the audit, it was concluded that a new seal oil system would be purchased and
installed prior to the turnaround. Since the original system was a combined lube /
seal system, an equalization line would be installed between the 2 reservoirs in the
event oil crossover at the seal / bearing interface would occur. Care was taken to
assure the volume of the new reservoir would not lower the oil level in the existing
reservoir.
Step Seven - Calculate new Process Vent Time
1.
2.
The audit revealed that remote operated block valves had to be added to the system
to reduce the volume of gas trapped in the compressor during emergency block in
and vent.
Calculations were made, taking into account the maximum 5 psi. / sec. vent rate to
avoid damage to the compressor o-rings. This required use of a cascade valve
system, with the second valve opening at a lower system pressure to minimize the
vent time.
Step Eight - Calculate Oil Volume available for Emergency Blackout Condition
1.
The volume available in the existing overhead tank during Emergency Blackout, for
the new conditions was only 20 seconds, where API requires 3 minutes, however the
exact requirement is calculated in Step Seven.
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SAUDI ARAMCO CONSULTING SERVICES
ROTATING EQUIPMENT / VIBRATION TECHNICAL EXCHANGE
2.
Step Nine -
One solution was to increase the size of the overhead tank, however the turnaround
time eliminated this as a possibility. A pressurized rundown tank, at grade level was
used to provide the additional flow during the Power Blackout Condition, (Figure
Fourteen). This required careful sizing of the new reservoir to assure it would not
overflow during rundown.
Calculate Compressor Endwall and Seal Component Pressure Drop on Oil
Supply Side
Due to an original concern with subsynchronous vibration, the plan was to run with
the existing breakdown seal clearances. However an audit of the endwall drops
indicated the unit could blow gas during the Power Blackout Condition.
1.
2.
Step Ten-
A detailed audit of the Compressor Rotor Dynamics, indicated reduced clearance
breakdown bushings could be installed. This minimized the flow during settleout and
alleviated the need to redrill the endwalls.
A backup plan was developed in parallel to the rotor dynamics study, to reduce the
endwall pressure drop. A detailed review of the endwall drawings indicated the
original design was set up to accept either forward or backpressure control. The
backup plan involved utilizing the backpressure porting in parallel with the supply
porting to reduce the pressure drop, in the supply porting.
Calculate Seal Component and Compressor Endwall Pressure Drop
on Oil Drain Side
1.
2.
An audit of the drain porting in the seal components indicated a velocity over 5
ft./sec. This also applied to the combined lube / seal oil return to the original single
reservoir. With the addition of the new reservoir, the return lines could be separated
and sized to 2 ft. / sec. to avoid and oil backup into the bearing housing drains.
Finally, the drain porting in the seal components was increased to lower the flow
velocity to 1 ft. / sec.
SUMMARY
During rerating of Centrifugal Compressors with oil type seals, performing the Ten Key Steps, is
critical to assuring the flawless execution and eventual commissioning / long term operation of the
Rerated Equipment. As new designs continue to progress to Dry Gas Seals, expertise in the O.E.M.
continues to fade. Audit Engineers should at a minimum review each of the areas discussed to assure
the new design is in compliance with the intent of the original equipment design.
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Author
Roy J. Salisbury is a graduate of the University of Pittsburgh, with a Bachelor of
Science Degree in Mechanical Engineering / Aerospace Option. He has over 25 years
experience in the installation, commissioning, and troubleshooting of Rotating
Equipment. He has previously worked as Field Service Manager at both Elliott-Ebara
Turbomachinery and Siemens Demag Delaval Turbomachinery.
He is currently Senior Rotating Equipment Engineer for ExxonMobil, out of their Fairfax
Virginia, Research and Engineering Facility. This group supports ExxonMobil's interest
in 42 refineries worldwide. His group is responsible for the selection, testing, installation,
commissioning, and troubleshooting of all rotating equipment in the Downstream
Organization.
Roy recently completed commissioning support of the Lubref II and Yanpet
Expansion Projects in Yanbu Al-Sinaiyah.
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